Power plants comprising hydraulic torque converters



Sept. 24,v 1963 K. G. AHLEN 3,104,560

POWER PLANTS COMPRISING HYDRAULIC TORQUE CONVERTERS Filed June 15, 19564 Sheets-Sheet l .la F157;! 1 1b 10,

1 d 44; f M E 6 0 a 1 5 by 6 "44 1 Jla. 4 a 42 m W a w I 60 5 my! I I rJ C 21 a Y "71 INVENT 0R BY W ATTORNEY Sept. 24, 1963 K. G. AHLEN 3,

POWER PLANTS COMPRISING HYDRAULIC TORQUE CONVERTERS Filed June 15, 19564 Sheets-Sheet 2 /)M a a) W s a a J I ll A I /d 15 INVENTOR M (9% M B Md M ATTORNEY K. G. AHLEN 3,104,560

POWER PLANTS COMPRISING HYDRAULIC TORQUE CONVERTERS Sept. 24, 1963 4Sheets-Sheet 3 Filed June 15, 1956 Sum wNm INVENT OR BY W MATTORNEY K.G. AHLEN 3,104,560

POWER PLANTS COMPRISING HYDRAULIC TORQUE CONVERTERS Sept. 24, 1963 4Sheets-Sheet 4 Filed June 15, 1956 INVENTOR BY w Mummy United StatesPatent 3,104,560 POWER PLANTS COP/WRISING HYDRAULIC IQRQUE CGNVERTERSKarl Gustav Ahln, Stockholm, Sweden, assignor to Svenska Rotor MaskinerAktiebolag, Nae-Era, Sweden, a corporation of Sweden Filed June 15,1956, Ser. No. 591,586 2 (Zlaims. (Ci. 74-688) This application is acontinuation-in-part with respect to my copending application Serial No.29,445, filed May 27, 1948 (now Patent No. 2,853,855, granted September30, 1958), and relates back thereto, so far as common subject matter isconcerned, to all dates and rights incident to the filing thereof and ofthe filing of the corresponding Swedish application Serial No. 4886/47filed May 28, 1947.

The present invention relates to power plants and has particularreference to power plants including hydraulic torque converters of thehydrodynamic kind. Still more particularly the invention relates topower plants in which such converters are combined with combustionengine prime movers.

The invention is particularly applicable for use in the propulsion ofvehicles and other forms of traction drives, and while its utility isnot limited to such use it will here inafter be discussed and itsadvantages pointed out in connection with automotive vehicle drivesbecause of its particular applicability to that kind of drive. 7

In vehicle drives employing variable speed power transmitters, it isordinarily desirable to provide a relatively high ratio of torquemultiplication under conditions of stall of the driven element, forproducing rapid initial acceleration and also in order to have hightractive effort available at low vehicle speeds. In hydraulic torqueconverters of the hydrodynamic type as heretofore developed, the desiredhigh starting or stall torque ratio has been obtained in either one oftwo ways, either by providing at least a three-stage turbine in thehydraulic circuit, through the use of which torque multiplication atstall of the order of 5 to 6 or more to 1 are obtained, or by combininga torque converter provided with a turbine having only one or two stageswith some form of auxiliary torque multiplying gearing for furtherincreasing the stall torque ratio under emergency or other conditionsrequiring a higher torque multiplication than a one or two stage torqueconverter is capable of providing, unless the converter is so geared tothe vehicle wheels as to materially reduce the maximum vehicle speedefficiently obtainable in hydraulic drive, as compared with usualstandards. Both of these expedients are relatively expensive eitherbecause of the cost of providing multiple stage converters with arelative- 1y large number of blade rings or because of the cost ofproviding an auxiliary gear.

Also, as heretofore constructed, torque converters usually have arelatively constant torque absorbing characteristic which results inpermitting the engine to rapidly speed up to maximum sped or a sped nearto the maximum speed when the throttle is opened under stall conditions.This results in relatively very quick application of the full power ofthe engine, at stall and very low tur bine shaft speeds, withoutresulting in a corresponding increase in secondary torque and also thistype of operation is undesirable in many other instances because of thenoise level resulting from rapid acceleration of an engine from idlingto full or near full speed while at stall and throughout theaccelerating period, and is also undesirable because of the poor fueleconomy resulting from such type of operation.

Modern combustion high speed engines have operating characteristicswhich are quite different from those of 3,184,588 Patented Sept. 24,1963 ice Typical of such engines are engines which produce maxi-' mumtorque at speeds approximating half the speed at which full power isdeveloped, the torque increasing from idling speed to maximum atsomewhere in the neighborhood of half speed and thereafter droppingmaterially with increase of speed to the speed at which full power isattained. In some instances the drop in torque can amount to as much as40% or more from half speed to full speed.

Bearing in mind the above characteristics of certain kinds of enginesand also having in mind the torque converting characteristics of theturbine type of hydrodynamic converter, it is a general object of thepresent invention to provide a new and improved form of hydrauliccircuit, which, particularly in combination with an engine havingvarying power and torque characteristics of the general nature statedabove, will provide improved overall tractive effort performance over awide speed range inclusive of the required high driven shaft torque atstall and with high peak efficiency, with a fewer number of stages ofblading than heretofore has been-required in a comparable installationand with torque absorbing characteristics which enable the fullestpossible use and advantage to be taken of the output torquecharacteristics and flexibility of the engine. A still further object ofthe invention is to provide a novel form of hydraulic torque converterwhich while having the characteristics noted above is capable of-operation in two diflerent ways in two different speed ranges of thedriven or output member, so that in the low speed range torquemultiplication equal to or exceeding that obtained with a larger numberof stages is secured without increase in the number of rows of bladesrequired, while in the high speed range of the driven member higherefiiciency is maintained and torque multiplication or increase isobtained up to a relatively high ratio of the speed of the driven memberto that of the driving member. Still another object is the provision ofa novel torque converter arrangement having the advantages abovediscussed which will further enable the hydraulic torque convertingmechanism to be used in alternation with a direct mechanical drive.

Still further objects of the invention and the advantages to be derivedfrom its use will best be understood from consideration of the ensuingportion of this specification in which, by way of example but withoutlimitation, different suitable embodiments of apparatus for carrying theinvention into effect are described in conjunction with the accompanyingdrawings forming a part hereof, in which:

FIG. 1 is a longitudinal central section of the hydraulic circuitportion of a torque converter embodying the principles of the invention;

FIG. 1a is a section taken on the line la-1a of FIG. 1 and includingalso a velocity diagram showing the nature of the flow of the fluidleaving the last turbine blades at stall;

FIG. 1b is a section taken on the line lblb of FIG. 1 and including avelocity diagram showing the flow of the fluid at the pump inlet atstall;

FIG. 1c is a flow diagram similar to the diagram of FIG. la but showingflow conditions at the changeover point;

FIG. 10! is a velocity diagram similar to the diagram form of converterembodying the principles of the invention;

FIG. 3 is a more or less conventionalized diagram iiustrative of certaintorque characteristics of converters embodying the invention;

FIG. 4 is a central longitudinal section of a converter embodying theinvention and incorporating a mechanical direct drive for use inalternation with converter drive;

FIG. 5 is a section on the line 5-S of FIG. 4;

FIG. 6 is a section similar to FIG. 4 showing another and similar formof converter with direct drive embodying the invention;

FIG. 7 is a section taken on the line 77 of FIG. 6; and

FIG. 8 is a diagram illustrative of the efficiency and torquecharacteristics of converters of the kind shown in FIGS. 4 and 6.

Referring now more particularly to FIG. 1 the mechanism illustratedcomprises a rotationally stationary casing 16 in which is rotatablymounted a pump wheel constituting the driving member of the apparatusand comprising a pump disc 24 carrying a ring of pump blades 28 andhaving a hollow hub or shaft part 24a adapted to be connected to thecrank shaft or other power output member of a prime mover. Also mountedfor rotation in the torus shaped chamber provided by casing 16 is thedriven or turbine member comprising the turbine disc 42 and hollowturbine hub or shaft part 44. The disc 42 carries two rows of turbineblades 34 and 36, the former being located radially outside the pumpblades 28 in the radial outflow part of the circuit while the latter arearranged in the radial inflow part of the circuit.

A ring of blades 46 situated in the circuit between blades 3-4 and 36 iscarried by disc 5t connected to or formed integrally with the hollow hubor shaft part 48. For the sake of convenience the blades 46 will bereferred to as reaction blades although as will hereinafter be explainedin certain of the constructions embodying the invention these blades mayconstitute rotating rather than stationary reaction blades, in whichcase they also partake of the nature of turbine blades since theytransmit power to the driven members when rotating in a direc tionopposite that of the pump blades.

Advantageously, but not necessarily, the turbine and reaction bladinghas the general form of profile with bluntly rounded inlet sectioncharacteristic of the blading disclosed in Lysholm US. Patent No.1,900,118 and as will be seen from FIGS. 1a and 1b the angles at whichthe blades are placed is such that when the hydraulic liquid iscirculated by the pump in the direction indicated by arrow 56, it willtend to rotate blades 34 and 36 in the same direction as the directionof rotation of the pump and will tend to turn the reaction blades 46 inthe opposite direction of rotation.

The general organization of the apparatus shown is known and is adaptedto be employed in conjunction with a mechanical direct drive in themanner illustrated in Lysholrn US. Patent 1,900,119, a direct driveshaft 8 passing through the hollow parts 24:: and 44 being adapted to beconnected in alternation directly with the power output element of theprime mover or with the pump member 24a by means of suitable clutchesand the turbine or driven member 44 being adapted to be connected to thedirect drive shaft 8 through the medium of an overrunning clutch. Sincethese structural arrangements are well known in the art (see Lysholm US.Patent No. 1,900,119) it is not necessary to illustrate and describethem herein for an understanding of the present invention. The reactionblades 46 may be either fixed against rotation by connecting the shaft48 to the stationary casing 16, or as will hereinafter be more fullyexplained, the shaft member 48 may be connected by means of suitablegearing to the driven shaft member 44 so that blades 46 '4 instead ofremaining stationary will rotate in a direction opposite to thedirection of rotation of blades 34 and 36.

An important departure from prior constructions embodied in the presentarrangement lies in the positions of the turbine and reaction bladingrelative to .the pump and particularly the relative radial positions ofthe last turbine stage and the pump, and as will be observed from FIG. 1the blading is so arranged that the last stage of turbine blading 36which rotates in the same direction as the pump discharges directly tothe inlet of the pump without there being interposed between theseblades any guide or reaction blades or turbine blading rotating inopposite direction. Also it will be observed that by placing thereaction blades and the last stage of turbine blading on the radialinflow side of the circuit opposite the side in which the pump islocated, and further by placing the first stage of turbine blading 34immediately outside the discharge edge of the pump blades, sufiicientroom is made available so that the outlet edges 11 of the turbine blades36 which discharge to the pump can be located at a radial distance fromthe axis of rotation that is sufliciently large without using too smallblade profile sizes.

The importance of the above described general arrangernent and relationof the blading lies in the fact that with such an arrangement thecharacteristics of the converter are made such that as the speed of thedriven or turbine member decreases relative to the pump speed towardstall, the torque absorbing capacity of the COD. verter rises relativelysharply, so that the load on the engine and consequently its speed evenif operating at full throttle is materialy reduced as the turbineapproaches stall condition. This provides a highly desirable operatingcharacteristic for the power plant as a whole as will hereinafter bemore fully explained.

The reason for the increasing torque absorbing characteristic with ablade arrangement such as that above described may best be understoodfrom a consideration of the following discussion of the factorsinvolved, with reference to the diagrams shown in FIGS. la-ld of thedrawings.

The hydraulic head produced by, a vane type impeller such as the pumpemployed in converters of the kind under consideration, which is ameasure of the input torque required to turn the pump, is expressed bythe following well established formula:

H is the pressure head developed by the pump;

up is the pump efficiency;

g is the gravity;

u is the peripheral velocity of the pump at the outlet;

u is the peripheral velocity of the pump at the inlet;

C is the absolute speed of the working fluid;

C is the projection of C on the tangent of the pump circle at theoutlet;

C is the projection of C on the tangent of the pump circle at the inlet.

From the formula it will be immediately apparent that the value of Hwill change with change in the value of the factor C presuming 11 and Qunchanged and in accordance with the present invention H is caused tomaterially decrease as the ratio of the secondary or turbine speed 11 tothe primary or pump speed n (12 /11 increases. The variation of thefactor u fl depends upon the variation of energy after the turbinedelivering fluid .to the pump.

If we presume a torque converter that has a certain inlet diameter ofthe pump we will get the following. results. Let us first consider thevelocity diagrams form-- ing parts of FIGS. 1:: and 1b which illustratethe how conditions obtaining at the outlet edge 12 of the last.

stage of turbine blading and the inlet flow conditions at the inlet edgea of the pump blading, when the pump is operating and the turbine isstationary or at stall, or in other words when the ratio n /n isinfinity. In the diagram of FIG. lzz the vector C indicates the absolutespeed and direction of the fluid leaving the blade 36 and also, sincethese blades are stationary, indicates as well the relative velocity wsince the peripheral velocity of the blades u equals zero. Theprojection of C on the tangent of the turbine outlet circle is, as shownin the diagram, in the opposite direction from the normal direction ofrotation of the blades, and is thus a minus value at stall. If thisminus value were to be substituted in the above equation (in other wordsif it were to be assumed that C were the same as C the expression inwhich this factor occurs would be negative and since the expression as awhole in the equation is negative the factor would in the equation bepositive, thus resulting in a higher value of H than would be the caseif c were positive.

It is, however, fundamental to hydraulic circuits of the kind underconsideration that the tangential component of free flow 0, increaseswith decrease in the radius of the circle to which the flow is tangent,and this increase is in direct proportion to the change in radius of thecircle. In this connection is to be noted that this change in tangentialvelocity is independent of the variation in the rate of circulation ofthe fluid in the circuit, that is, the quantity of fluid circulatedthrough the blade per unit time. The effect of this characteristic isbest illustrated by the diagram associated with FIG. lb whichillustrates the inlet flow conditions at the pump with the arrangementshown when the turbine is at stall. In this diagram the vector 11,,indicates the peripheral speed of the pump blading while the vector Crepresents the tangential velocity component of the fluid entering theblading. By comparison of the diagrams of FIGS. la and lb it will beseen that the vector C is substantially greater than the vector C Withthe blading laid out as shown in FIG. 1 the radius of the edges h isapproximately 30% greater than the radius of the edges a andconsequently vector C is approximately 30% greater than the vector C Thesense of the vector C is still negative and if substituted in theequation results in a substantially higher value of H than if the vectorC were substituted in the formula, which would be the case if the radiiof edges a and h were equal. Consequently by placing the outlet edges ofthe last turbine stage, which rotates in the same direction as the pumpand delivers directly to the pump, at a substantially greater radiusfrom the axis of rotation than the inlet edges or the pump blades,substantially higher torque absorbing characteristics are imparted tothe pump at stall than would otherwise be the case.

While the above stated condition at stall is for the purposes of thepresent invention highly desirable and may be varied in degree to suitindividual conditions by choosing the proper relation between the radiito the edges a and h, this condition would not be desirable if it wereto hold good as a'more or less constant relationship over the normalspeed range of operation of the turbine. However, such constancy of therelationship does not obtain as may be explained by reference to FIGS.and 1d. 'lihese diagrams are representative of conditions which in atypical design may obtain when the speed ratio 11 /12 is greater than0.5, which is representative of the higher portion of the normal speedrange of the turbine with respect to pump speed, converters of the kindunder consideration ordinarily reaching a condition of operation atwhich the output torque drops to a value equal to the input torque whenthe speed ratio fig/Il is of the order from 0.6 to 0.8.

In the diagram of FIG. 10 the peripheral velocity of the turbine bladesis represented by vector u The relative velocity of the fluid leavingthe turbine blades is shown by vector w and due to the peripheralvelocity a the absolute velocity of the fluid leaving the blades is asshown by vector C,,. The tangential projection of this Velocity is shownby vector C and it will be observed that the direction of thistangential component of velocity is the same as the direction ofrotation of the blades, so that the sense or sign of this faflor ispositive, rather than negative at stall.

If this positive value is substituted in the above formula the factor inthe equation of which it is a part is positive and is consequentlysubtracted rather than added to the remaining factors in the equation.'The result of this is that the hydraulic head produced by the pump isless than at the condition of stall.

If we now consider the diagram of FIG. 1d the peripheral velocity at theinlet of the pump is shown by vector u,,, the relative inlet velocity ofthe fluid to the pump by vector W and the absolute velocity of theliquid at the pump inlet by the vector C The tangential component of theabsolute velocity is shown by the vector 0,, and again as in the case ofthe condition at stall, due to the difference in radii between the bladeedges 11 and a, the tangential component C is larger than the componentCm, the diiference in magnitude being proportional to the ratio of theradii a and h. It follows then that if we substitute the larger plusvalue of C in the formula, the result is a further decreased value of HAgain referring to the four diagrams of FIGS. la-ld, we find that theICSlllllI of moving the outlet edge of the last stage of turbinebladin-g to a greater radius than the inlet edge of the pump blading hasthe effect of materially increasing the torque absorbing characteristicsof the pump as the turbine speed is decreased from its normal operatingrange to stall, as compared with a construction in which these two setsof blade edges are on the same or nearly the same radii or with theturbine blade outlet edges on a shorter radius than the pump inletedges. If for example we assume that the edges a and h are on the sameradius, the difference in the value of H as between stall and a value of0.5 or above for the speed ratio n /n is represented by the change inthe values of the vectors C and c which here are equal, with the valueof H increasing but little as the turbine speed decreases from a given11 /11 ratio to stall.

On the other hand, with the present arrangement this variation in thevalue of H as the turbine speed decreases toward stall is represented bythe change in value of the product of the vectors C hXu and as will beevident from the preceding discussion this represents a very materialincrease in the amount by which the pump head is increased at stall byplacing the last turbine outlet edges materially further from the axisof rotation than the inlet edges of the pump blades.

It is an established tfact that the rate of circulation of the Workingfluid in the circuit is a function of the hydraulic bead H developed bythe pump, the rate increasing with increase in the hydraulic head.Consequently the present construction which tends to rapidly increasethe value of H as the turbine approaches stall also causes the rate ofcirculation to increase, with the net result that the torque absorbingcharacteristic of the converter is rapidly increased as the turbinespeed decreases. The desirability of this characteristic, particularlywhen combined with certain obtainable torque output character'- isticsof internal combustion engines, will later be shown, but beforeconsidering that phase of the invention a further characteristic of thenature of the increase in torque absorption as stall condition isapproached, as obtained by the present invention, must be taken intoconsideration.

As has previously been noted, it is a basic characteristic of convertersof the kind under consideration that the value of the secondary oroutput torque falls as the speed of the turbine member increases fromstall, to a value equal to the input torque before the speed of theturbine member reaches the speed of the pump, the value of the ratio 11/11 at which the torque ratio becomes 1:1 being ordinarily somewhere inthe range of from 0.6 to 0.8. The speed ratio 11 /11 at which the torqueratio becomes 1:1 is ordinarily referred to as the shift or changeoverpoint since when the torque ratio falls to 1:1 value some form of driveother than continued drive through the torque converter should beemployed if a still higher rate of driven shaft speed relative to pumpor engine speed is desired. It is highly desirable, for reasons whichwill later :be discussed, to provide for the purposes of the presentinvention a relatively high ratio between the values of the torqueabsorbing characteristic of the converter at the shift point and thetorque absorbing value at stall. In many instances it may be desirableto provide a construction in which the conven ter will absorb five tosix or more times as much torque at stall as at the shift-over point,and in some special cases it may be desirable to have the torqueabsorption at stall as much as ten times the torque absorption at theshift-over point. q

We have previously seen how in accordance with the present concepttorque absorption at stall as compared with torque absorption at or nearthe shift-over point can be materially increased by the relative radialpositions of certain of the blade edges. I have found that there is acertain necessary relationship between the radius of the last stageturbine outlet edges and the radius of the pump inlet edges which mustbe employed if the desired results are to be obtained. The nature ofthis relationship may best be understood by referring to FIG. la inwhich the ratio of the input torque at stall (M stall) over the inputtorque at shift (M shift) is plotted as ordinates, and the ratio of theradius of the last stage turbine blade outlet edge (r over the radius ofthe pump blade inlet edges (r,,) is plot-ted as abscissa.

In the diagram the curve M represents the ratio of the inlet torques atstall and shift, and as will be seen from the nature of the curve theincrease in the value of the ratio of the input torque at stall ascompared with input torque at shift does not increase as a straight linefunction with increase in the radius of the turbine blade outlet edgesas compared with the radius of the pump inlet edges. On the contrary,the more the latter ratio is increased the more rapidly the torque ratioincreases and as will further be seen from the curve, placing the lastturbine stage blades so that their outlet edges are at a radius Withinthe range of not more than approximately 15% greater than the radius ofthe pump inlet edges produces small appreciable effect in increasing theratio of the input torques at stall and shift. As shown by curve M,which is based upon test results, a value of approximately 2.5 for thetorque ratio M stall/M shift for a given converter was obtainable with ablade arrangement in which the turbine outlet edges 11 and the pumpinlet edges a were at approximately the same radius. In such aconverter, shifting of the turbine blade edges 11 to a radiusapproximately 10% greater than the radius of the pump inlet edges 0resulted in very little increase of the value of M stall/M shift, to avalue a little over 3. On the other hand, however, a change of positionof the turbine outlet edges h to a position having a radius 50% greaterthan the pump inlet edges a resulted in an increase in the value of Mstall/M shift to approximately 18, or in other words an increase in thevalue of the ratio of approximately seven times. A value of the ratio Mstall/M shift as high as 18 is one that usually will not be required.Other design factors make it ordinarily more diificult to properlydesign a converter with a ratio of r /r. of 50% or more than to designan equivalent converter with a smaller value of this ratio. Consequentlyin order to obtain the maximum benefit from the present invention, it ispreferable to so construct the blading that the ratio r /r is within arange of which the lower limit is approximately 1.15 and the upper limita ratio determined by practical requirements but ordinarily not overapproximately 1.5.

However, the pump inlet does not represent a constant value and furtherthe variation of the inlet radius of the pump has no influence on thesloping of the torque curve, which as previously mentioned is dependingon the variation in energy content between stalling and shifting pointof the circulating fluid when this one is leaving the turbine in frontof the pump.

As a fixed unchangeable factor for the relation to the outlet radius ofthe last turbine row we have chosen the outer diameter of the hydrauliccircuit, which is an internationally used measure of the size of thesystem. Thereby we have found that values for the ratio of M to M shiftwill vary between the limits 4 0.4+22.5: and 1.6+22.5

However, a z1 =r -w, where w is the angular speed of the pump. The lastfactor thus will become In the space bet-ween the last turbine and thepump, however, the relation n -C =r -C prevails. Thus the formula may bechanged to If we compare this formula with the previous one we find thatthe factor LL -C has been transformed to in which we presume that thefactor a: is constant.

We thereby have got a term for H in which r alone is included and not aspreviously in relation to the inlet radius of the pump. We thereforehave considered it more justified to refer r to the outer radius r ofthe hydraulic circuit.

At stalling r -C is negative, as has been previously shown, which willincrease the value of. H and thereby M stall. This increase will begreater at increased r At the shifting point r -C is positive, alsopreviously shown, and will decrease the value of H and thereby M shift,whereby the decrease will be greater at increased r From this it willfollow that M stall/M shift increases if r increases. This means thatfor a constant input torque the speed of the engine will be lower atstall than at shifting point and so much lower as the square root of therelation between torque absorption at stall to the torque absorption atshifting point. Other factors than those mentioned which have anyinfluence have, however, been proved to be so small compared with thevariation of r that the same are included Within the limits for thecurves defined by the formulas.

In connection with the foregoing discussion it is further desirable topoint out that change in the outlet angle of the turbine blading,particularly that of the last turbine stage, also influences, thecharacteristics of the input torque absorbing characteristics of theconverter. If the outlet angle of the last stage turbine blades isincreased the effect will tend to decrease the value of the torqueabsorption of the converter over the entire range from stall to shiftpoint, and vice versa. However, the effect of vary- 9 a ing the outletangle of the last turbine stage blading is less for a relatively highvalue of the r /n than for smaller values of that ratio.

With blade arrangements embodying the present invention a widervariation in the outlet angle of the last stage turbine blading isavailable to the designer, to meet specific conditions than hasheretofore been the case. referably this angle will not be less thanapproximately 20 as a lower limit, but may in certain more or lessextreme cases be made as high as 90 or even greater, the latter caseproviding what may be considered a negative outlet angle. Usually,however, the lower limit will not be less than approximately 35 and theupper limit will not exceed approximately 55.

With a converter having the characteristics provided by the principlesabove discussed in mind we will now consider the nature of the improvedresults obtained in a power plant as a whole which combines such aconverter with an internal combustion engine of the kind having arelatively wide speed range and further having an output torquecharacteristic which provides a materially decreasing torque from amaximum value which occurs at approximately the mid-speed range to asubstantially lower value at both low speed and the speed at whichmaximum power is developed.

In the preceding discussions the various factors and characteristicshave been considered on the basis of operation of the converter atconstant input shaft speed, but as will have already been understood theobject of the invention is to provide a converter and power unit whichwill result in actual vehicle operation which will make use of highlyvariable input shaft speeds and in connection with the followingdiscussion concerning the cooperative relation between a converterembodying the present invention and a variable speed, variable torque,internal combustion engine, it must be borne in mind that, otherconditions being equal, the torque absorbing characteristic of aconverter of the kind under consideration is such that the torqueabsorption for any given ratio n /n varies substantially as the squareof the input shaft speed.

The invention is applicable to many specific kinds of converters and inFIG. 2 there is shown a blade system incorportaed in a converter havinga rotating casing rather than a stationary casing with a through directdrive shaft as in FIG. 1. As shown in FIG. 2, the primary or drivingmember 24 is constituted by a rotating casing adapted to be connected toand driven by the engine (not shown). The casing carries the pump orimpeller blades 28. The turbine disc 42 carries the two stages ofturbine blades 34 and 36 while the reaction blades 46 are carried by thereaction disc 50 as previously described. In addition to embodying arotating casing, the present example differs from the preceding designin that the first stage turbine blades 34 are located on the radialinflow side of the circuit so that longer pump blades are provided.

In FIG. 3 a more or less conventionalized diagram indicates thecharacteristics of the converter shown as compared with earlier knowntypes of converters, in terms of the in ut-output speed ratio n /n Inthis diagram the solid lines M and n indicate respectively the inputtorque absorbing characteristics and the resultant input speedcharacteristics of the present converter while the dotted lines M and 1indicate the corresponding characteristics of what may be termed aconventional prior form of converter.

In the description and discussion with reference to the precedingembodiments the nature and characteristics of the invention have beenconsidered without regard to the factor of whether or not the reactionblading 46 is restrained during converter operation against rotation ina direction opposite that of the pump and turbine or is incorporated inthe apparatus so that when the converter is operating this reactionblading rotates in the opposite or counter direction and becomes ineffect moving rather than stationary reaction blading and also through agear connection operates :to transmit torque to the driven memher andthus also may be considered as counter-rotating turbine blading.

The use of counter-rotating turbine blading is broadly known and forconvenience converters embodying this construction will hereinafter bereferred to as double rotation converters as distinguished from singlerotation converters in which the reaction blading is restrained againstcounter-rotation.

The present invention is particularly advantageous when incorporated indouble rotation converters, for reasons hereinafter to be pointed out,and by way of example there is shown in FIGS. 4 and 5 an embodiment ofthe invention in the form of a double rotation converter combined with adirect drive.

Referring now to FIGS. 4 and 5, the hydraulic circuit illustrated isembodied in a structure of the rotating casing type in which therotating casing 24, which is driven from the engine flywheel 12 throughthe spline or tooth connection 32, carries the pump blades 28. Theturbine member 44, which is carried by bearings 58 and at), is providedwith a wheel portion 42 which carries the two rows of turbine blades 34and 36. Between these rows of blades, the row of reaction blades 46 iscarried by disc 5% forming a part of the reaction member having a hollowshaft portion 48 to which is keyed a hollow shaft extension 62. Betweenthe turbine or driven shaft '44- and the extension 62 of the reactionmember there is located a free-wheel clutch lit) so arranged that if thereaction member tends to turn in the same direction as the driven shaftmember 4-4- the clutch will engage to thereby prevent the element 62from over-running element 4%, while freely permitting relative rotationbetween the parts in opposite directions.

Member 62 further is provided with a sun gear 366 mes-hing with planetgears 374} which in turn mesh with an internal ring gear carried by amember 368 keyed to the driven member 44. The planets 376 are carried bya suitable planet carrier 376 between which and the stationary casing 16there is located a free-wheel brake 373 so arranged as to engage toprevent the planet carricr from rotating in a direction opposite that ofthe pump and turbine members and to permit the planet carrier to rotatefreely in the same direction as these mem bers.

A multiple disc friction clutch is interposed between the rotatingcasing 24 and the reaction member. This clutch comprises a number ofaxially movable clutch plates 27d keyed or splined on the extension 1212of the rotating casing and a number of interleaved clutch plates 254axially keyed or splined on a bell-shaped extension 252 of the reactionmember 4a. The latter also carries an axially stationary backing plate256 and an axially movable clutch actuating plate or piston located in asuitable annular recess in the member 252 and adapted to be moved toengage the clutch by hydraulic pressure admitted to or released from thechamber 264 under the control of a valve member 356'.

The valve member 35% is actuated to cause engagement or disengagement ofthe clutch through the medium of a pivoted fork 354 (see FIG. 5) theaction of which is controlled by a system of hydraulically actuatedservomotors 352 and 362 to whichpressure fiuid is admitted 'or releasedunder the control of the axially movable valve 396. Fluid under pressurefor operating these servomotors and also for maintaining desired basichydraulic pressure in the converter circuit is supplied by means of thegear pump 38% driven from the extension 122 of the rotating casingthrough the medium of the intermediate gear 384 meshing with gears 386on the part 122 and gear 382. on the pump shaft.

The outer circumference of the extension 352 on the reaction memberprovides a brake drum surface adapted to be engaged by a band brake 35$for holding the reaction member rotationally stationary. This brakethrough the medium of the band 360 and actuating lever 364 is engaged byupward movement of the piston of the servomotor 362 and released by theaction of the spring forming a part of that device. A spring arrangement410 is provided for retracting the brake band to prevent dragging whenthe latter is released.

The action of the servomotor arrangement is such that when the valve 396is moved to the position a, pressure fluid is not admitted to the clutch250' so that the latter is released to disconnect the rotating casingfrom the reaction member. Also in this position of the control valve thebrake band'360 is released to permit the reaction member to rotatefreely in either direction insofar as the brake is concerned. Underthese conditions, it will be seen by reference to FIGS. la and 1b thatthe reaction blades will rotate in opposite direction with respect tothe turbine blades 34 and 36, it being noted that in the presentconstruction the first turbine stage blades are located on the radialinflow side of the circuit rather than radially outside the pump blades28, as in the blade arrangement shown in FIGS. 1 and 2. This conditionprovides for double rotation operation of the converter, the torque fromthe reversely rotating reaction blades (which under this condition arein elfect also turbine blades) being transmitted through the reactionelement 62, gears 366, 370 and 368 to the driven member 44. The natureof the torque developed is such that it will tend to cause the planetcarrier 376 to rotate in a direction opposite that of the driven member,but this is prevented by the action of the free-wheel brake 3 78. If thevalve 396 is shifted to position b the resultant action of theservomotor system will be to cause the band brake 358 to engage brakedrum surface on the extension 252- of the reaction member and hold thelatter member against ro tation, while at the same time still notadmitting actl1- ating fluid to the clutch 250.. Under this conditionthe blades 46 become stationary reaction blades and the converteroperates as a single rotation converter. With the reaction member lockedagainst rotation, the sun gear 366 of the planetary gearing isstationary while the ring gear connected to the driven member continuesto rotate in forward direction. This, of course, requires that theplanet carrier 368 also rotate in forward direction, and this action ispermitted by the free-wheeling action of the free-wheel brake 378between the carrier and the stationary casing.

If the control valve 396 is moved to position the servomotors act toagain release the brake 358 and also act to actuate the member 354 so asto move the valve 350 to the left as seen in FIG. 4 and thereby admitpressure fluid to the chamber 264 behind the clutch actuating plate 258so that the clutch 250 is engaged. This serves to mechanically connectthe rotating casing directly with the reaction member which in turntransmits power mechanically to the driven member 44 through the mediumof the free-wheel clutch 110. Since under this condition the reactionand driven members travel at the same speed in forward direction thereis no relative movement between the gears of the planetary gear systemwhich is permitted to rotate in forward direction as a unit through theoverrunning action of the free-wheel brake 378. Also since the bandbrake 358 is released the clutch assembly is also free to rotate.

So far as the present invention is. concerned any desired specific formof clutch and the manner of its actuation may be employed andconsequently the servomotor system for actuating the clutch is notdescribed herein in greater detail than necessary to understand thefunctioning of the transmission illustrated in order to secure doublerotation, single rotation and direct drive. The details of the clutchand brake actuating mechanism shown form a portion of the claimedsubject matter of my US. Patent No. 2,719,616, granted October 4, 1955,which matured 12 from application Serial No. 29,446, filed May 27, 1948,in which this portion of the apparatus is described in greater detailand to which reference may be had.

Insofar as the specific hydraulic circuit shown in FIG. 4 is concerned,it will be evident that this embodies the principles previouslydiscussed in connection with FIGS. 1 and 2, and it will be evident thatthe specific structure shown in FIG. 2 may readily be incorporated inthe organization shown in the present figure.

In FIGS. 6 and 7 another example of double rotation converter embodyingthe principles of the invention is illustrated. The construction of thisembodiment is similar in all respects to that shown in FIG. 4- exceptfor the arrangement of the planetary gearing connecting the reaction andturbine members and need not again be described in detail. In thepresent arrangement the extension 252 of the reaction members carries anaxially extending annular flange providing a ring gear 504- meshing withpinions 500 which are carried on pins 502 extending from the stationarycasing 16. Pinions 500 mesh with a sun gear 512 between which and thedriven member 44 there is located the free-wheel clutch 514. Thefreewheel clutch 514' is arranged to permit the gear 512. to rotateoppositely relative to the driven member 44 and to engage to preventgear 512 from overrunning the driven member 44 in the same direction.

In the operation of this form of the transmission the position a ofvalve 596 results in release of the clutch 250 and the band brake 258 topermit the reaction blades to rotate in reverse or counter-direction.Torque from these blades is transmitted through the reaction member togear 504 and through the medium of the pinions 500 the direction of thedrive is reversed so that torque is applied through clutch 514 inforward direction to the driven member, double rotation operation thusbeing eflfected. In position b of valve 596 the band brake 358 engagesmember 252 to lock the reaction member against rotation and securesingle rotation operation of the con verter. When the reaction member islocked against rotation this also prevents rotation about theirindividual axes of the pinions 500 and likewise prevents rotation of thesun gear 512. This gear can, however, remain stationary while the drivenmember 44' is turned in forward direction due to overrunning action ofthe clutch 514.

Movement of the control valve 596 to position 0 again releases the bandbrake 358 and admits pressure fluid to clutch 250 to engage the latter.Under this condition the reaction member which is now mechanicallyconnected to the rotating casing transmits power from the latterdirectly to the driven shaft through the medium of the overrunningclutch which is arranged so as to engage whenever the reaction membertends to overrun the driven member in the same direction. Under thiscondition of drive the ring gear 504 on the reaction member rotates inforward direction and through the action of the pinions 500 causes thesun gear 502 to rotate in a direction opposite that of the drivenmember, this being permitted by the action of the overrunning clutch514.

As previously noted, the principles of the invention are particularlyadvantageous when embodied in a converter capable of double rotationoperation. The reason for this is that inherently it is possible toprovide a higher ratio of torque multiplication (Mg/M1) with a doublerotation converter than with a single rotation convertcr,.

other things being equal. Briefly, the reason for this is that in anytorque converter the secondary torque M must always equal the sum of theprimary torque M and the reaction torque R transmitted to the stationarycasing or other stationary abutment. In the case of a single rotationconverter the reaction torque R transmitted to the stationary abutmentis equal to the torque R. which is hydraulically applied by the workingfluid to the reaction blading. In the case of a double rotationconverter, however, the torque R transmitted to the stationary abutmentis equal to the torque R plus the value of the torque R multiplied bythe gear ratio (which we will call k). Thus, other things being equal,the value of M at stall for example will be larger with a doublerotation converter than with a single rotation converter. As an exampleof the difference in stall torque ratios obtainable with the two typesof converters it may be pointed out that with present three-stage singlerotation torque converter having acceptable efiiciency and othercharacteristics, the maximum stall torque ratio usually obtained is ofthe order of from five to one to six to one. Theoretically, with adouble rotation converter an extremely high stall torque ratio isobtainable if other desirable characteristics are sacrificed, but withdesigns which provide satisfactory efficiency and range characteristics,stall torque ratios as high as twelve to one are readily obtainable in atwo-stage converter.

Bearing in mind the high stall torque ratio that can be obtained with adouble rotation converter, the particular advantage of a converter ofthis type embodying the principles of the present invention is found inthe fact that with a high stall torque ratio built into the converterand a high value of M stall/M shift also built into the converter by arelatively high ratio of r /r or r /r the latter characteristics can beutilized to pull the engine speed at stall down to comparatively a verylow value, even approaching a speed of the order of ordinary idlingspeed in extreme cases. By way of example, let it be assumed that thenormal full speed of the engine is 3600 rpm. and that the converter isbuilt with a ratio M stall/M shift which will reduce the full throttlespeed of the engine at stall to 1200 rpm. This reduction in the speed ofthe engine will obviously reduce its power at stall as compared with itsfull speed power, but owing to the fact that the double rotationconverter can readily provide a torque multiplication as high as ten ortwelve, or even more, to one, the desired tractive effort at stall canbe obtained even with the engine operating at comparatively very lowspeed. Thus with this arrangement a single more or less standardizedconverter is adapted to be combined with engines of widely varyingcharacteristics to give desired tractive effort results, since by thesimple expedient of adjusting the gear ratio between the double rotationparts and also adjusting the ratio r /r or r /r the converter can begiven torque absorbing characteristics and stall torque characteristicswhich will pull any given engine down at stall to any specific desiredspeed to give best fuel economy and other operating characteristics,while at the same time providing satisfactory tractive elfortcharacteristics for the vehicle.

As has been noted above, the difierence between the arrangements shownin FIGS. 4 and 6 is essentially in the gearing between the reaction anddriven members. In the example shown in FIG. 4 the reaction blading 46is connected to the sun gear 366 which is approximately half thediameter of the ring gear 368 connected to the driven or turbine member44. Thus the ratio k of the gearing is 2.0, so that the value of thereaction torque and consequently the secondary torque M is higher thanif the blades 46 were not so connected as to give a multiplied torque.In the case of the example shown in FIG. 6, the reaction blades 46 areconnected with the ring gear 564, which is substantially twice thediameter of the sun gear 512 connected to the turbine member. In thisinstance the gear ratio is 0.5 and while the torque actually appliedhydraulically to the blade 46 in this example is reduced in value beforebeing applied to the reaction and driven elements, the total torquemultiplication is still greater than for a single rotation convertersince the value of R k is a factor additive to the factors determinativeof the torque multiplication in a single rotation converter.

Thus in the example of gearing shown in FIG. 4, stall torquemultiplication will be greater, other things being equal, than with thearrangement shown in FIG. 6.

On the other hand, other factors than the stall torque ratio areinvolved by varying the gear ratio k. In the case of the construction ofFIG. 4 the gear ratio results in the reaction blading 46 rotating twiceas fast in counterdirection as the turbine blading 34, 36 rotates inforward direction and in this case the efiiciency rises from stallcomparatively rapidly as the value of n /n increases. Likewise theefiiciency of the converter in double rotation operation reaches itspeak and falls relatively rapidly at a comparatively low value of n /nthus making the desirable shift point from double rotation drive tosingle rotation drive at a comparatively low value of VIZ/I11 which isusually indicative of a relatively low vehicle speed.

In the case of the gearing shown in FIG. 6, the reaction blading indouble rotation operation turns in counterdirection at only half thespeed of the turbine blading 34, 36. In this case converter efiiciencyrises more slowly with increase in valve of n /n from stall and peak ef--ficiency in double rotation operation is achieved at a higher value offig/I'l so that the proper shift point to single rotation operation willoccur at a higher value of "2/111 and a higher vehicle speed than withthe construction shown in FIG. 4, other factors being equal.

Thus it will be seen that choice of the specific gear ratio to beemployed will be dictated by considerations such as the maximum stalltorque ratio desired and the range of the efficiency curve in terms ofva-tiation in the value of n 11 in individual cases.

In FIG. 8 there is shown more or less conventionally the nature of theefiiciency curves obtained with double rotation converters such as shownin FIGS. 4 and 6. In this figure the secondary torque is indicated bythe curve M the efficiency in double rotation operation by a, theefliciency in single rotation operation by b, the efficiency in directdrive by c, and the primary speed characteristics by n If for purposesof comparison the curves shown in FIG. 8 are assumed to berepresentative of the action obtained with a converter of the kind shownin FIG. 4, the effect of changing the gearing to one such as shown inFIG. 6 would be to decrease the value of M at stall and move peak of thecurve a to the right as seen in FIG. 8 so as to place the peakefliciency point of this curve at a higher value of "2/ n While forpurposes of illustrating the principles of the invention as exemplifiedin both single and double rotation types of converters, specific formsof double rotation converters having rotating casings and also havingmeans for providing single rotation operation and direct drive which isachieved in a particular fashion, the invention is not limited to suchspecific features of construction since it will be apparent that thestationary casing type of hydraulic system with a through shaft form ofdirect drive such as is illustrated in FIG. 1 may readily be combinedwith gearing of the kind shown in FIGS. 4 and 6 to provide both doubleand single rotation operation. Also it will be evident that in caseswhere desired the feature of single rota-tion operation or the directdrive feature or both may be omitted. Furthermore, while in theinterests of simplicity and minimum cost, two-stage converters of thekind herein illustrated are preferable and will in most instances meetthe requirements for traction drives, the principles of the inventionare as readily applicable to converters provided with a larger number ofstages of turbine and reaction blading.

From the foregoing it will be evident that the invention may be appliedin many specific difierent mechanical embodiments and that if desiredcertain features may be employed to the exclusion of others. Theinvention is accordingly to be considered as in no wise limited in itsscope to the forms of construction herein disclosed by Way of examplebut is to be considered as embracing in its scope all structures fallingwithin the purview of the appended claims.

What I claim is:

1. A hydrodynamic torque converter comprising within a housing a pumpmember having pump blades carried thereby, a turbine member havingturbine blades carried thereby, and a reaction member having reactionblades carried thereby, said housing having a Working chamber providinga hydraulic circuit for said blades, said turbine member being arrangedto be rotated by the working fluid in the same direction as the pumpmember and comprising a row of blades discharging working fluid directlyand unobstructedly to the pump blades, the ratio of the radial distancefrom the axis of rotation to the outlet edges of the blades of said rowof turbine blades to the outer radius of the working chamber beingwithin the range from approximately 0.50 to approximately 0.66, saidreaction member being arranged to be rotated by the working fluid incounter direction to the pump and turbine members, means fortransmitting the torque from the counter rotating reaction member inforward direction to the turbine member, means for selectively holdingsaid reaction member rotationally stationary or releasing said reactionmember to permit the rotation thereof in either direction, and meansproviding a direct driving connection between said pump and said turbinemembers.

2. A hydrodynamic torque converter comprising within a housing a pumpmember having pump blades carried thereby, a turbine member havingblades carried thereby, and a reaction member having reaction bladescarried thereby, said housing having a working chamber providing ahydraulic circuit for said blades, the ratio of the radial distance fromthe axis of rotation to the outlet edges of the blades of said row ofturbine blades to the outer radius of said working chamber being withinthe range from approximately 0.50 to approximately 0:65, said turbinemember being arranged to be rotated by the working fluid in the samedirection as the pump member, said reaction member being arranged to berotated by the working fluid in a direction counter-to that of the pumpand turbine members, means for transmit ting the torque from the counterrotating reaction member in forward direction to the turbine member,said turbine member comprising a row ofblades arranged to receiveworking (fluid directly and unobstructedly from the pump blades, saidreaction member comprising a row of reaction blades located in saidcircuit to be traversed by the working fluid afiter the dischargethereof from the turbine blade row receiving fluid from the pump bladesand prior to its entry to the turbine blade row discharging fluid to thepump blades, said means for transmitting torque from the counterrotating reaction blading to the turbine blading comprising gearinghaving a part for transmitting reaction torque to a rotationallystationary element, means for selectively holding said reaction memberrotationally stationary'or releasing it to rotate in either directionand automatically releasable means associated With said gearing -forpermitting said reaction member to rotate in the same direction as saidpump and said turbine members.

References Cited in the file of this patent UNITED STATES PATENTS1,900,120 L-ysholm et al. Mar. 7, 1933 2,005,444 Weiss June 18, 19352,041,189 Rabe May 26, 193-6 2,349,350 Jandasek May 23, 1944 2,441,818Jandasek May 18, 1948 2,853,855 Ahlen Sept. 30, 1958

1. A HYDRODYNAMIC TORQUE CONVERTER COMPRISING WITHIN A HOUSING A PUMPMEMBER HAVING PUMP BLADES CARRIED THEREBY, A TURBINE MEMBER HAVINGTURBINE BLADES CARRIED THEREBY, AND A REACTION MEMBER HAVING REACTIONBLADES CARRIED THEREBY, SAID HOUSING HAVING A WORKING CHAMBER PROVIDINGA HYDRAULIC CIRCUIT FOR SAID BLADES, SAID TURBINE MEMBER BEING ARRANGEDTO BE ROTATED BY THE WORKING FLUID IN THE SAME DIRECTION AS THE PUMPMEMBER AND COMPRISING A ROW OF BLADES DISCHARGING WORKING FLUID DIRECTLYAND UNOBSTRUCTEDLY TO THE PUMP BLADES, THE RATIO OF THE RADIAL DISTANCEFROM THE AXIS OF ROTATION TO THE OUTLET EDGES OF THE BLADES OF SAID ROWOF TURBINE BLADES TO THE OUTER RADIUS OF THE WORKING CHAMBER BE-